Hydraulic motor



1937" E. K BENEDEK v 2 HYDRAULIC MOTOR Filed May 10, 1934 6 SheetsSheet i f] 1 vu onion :ELEK KEENEI: EI 1 E. K. BENEDEK HYDRAULIC MOTOR Dec. 7, 1937.

Filed May 10, 1934 s Sheets-Sheet 2 7, 1937- E. K. BENEDEK 2,101,829

HYDRAULIC MOTOR Filed May 10, 1934 6 Sheets-Sheet 3 EEK K- :EJENEDEK Dec. 7, 1937.

E. K BENEDEK 2,101,829

HYDRAULIC MOTOR Filed May 10, 1934 6 Sheets-Sheet 4 3mm :ELEK K EENEDEK E. K. BEN EDEK HYDRAULIC MOTOR Dec. 7, 1937.

6 Sheets-Sheet 5 Filed May 10, 1934 ELEK KEENE EEK- MFM Dec. 7, 1937 E. K. BEN EDEK HYDRAULI C MOTOR Filed May 10, 1934 6 Sheet s-Sheet 6 (NVEIQKJR.

ELEK KiEJENEBEK.

' ATmRNEYJ.

Patented Dec. 7, 1937 UNITED STATES PATENT OFFICE 7 Claim.

This inventionrelates to improvements in hydraulic motors of the radial piston type and its principal object is to provide a highly eflicient, hydrostatically and mechanically balanced, compact motor unit having starting and operating.

torque capabilities'comparable to the more efficient types of sions.

A further object is to provide a hydraulic rotary 10 motor of the radial piston type wherein the hy-v drostatic pressures are balanced about the center of relative rotation of the main driving member of the motor.

A further object is to provide supporting bearl5 ings for the piston carrying rotor so arranged with referenceto thei cooperating parts that all the relatively rotating working parts are maintained in accurate alignment and in undistorted condition during varying conditions of powerinput and output.

Still another object is to provide a hydraulic motor of the radial piston type wherein viscosity losses and flow restriction losses, particularlyat the valves, are reduced substantially to a rninimum. i

Still another object is to provide in a hydraulic motor of the type referred to a more compact valve arrangement for feeding operating fluid medium to the piston cylinders and discharging said medium therefrom.

Still another object is to provide, in a hydraulic motor of the type referred to, a compact and balanced arrangement of pistons, operating cams,

bearings and driving shaft for the piston carrying rotor thereby reducing as much as possible the tendency for the piston carrying rotor to tend to become disaligned with the relatively rotating parts due to unbalanced mechanical stresses.

Another object is to provide a hydraulic motor 40 of the radial piston type wherein the mass of a the rotor is always in dynamic balance about the axis of rotation of the rotor.

Other objects include the provision of various inter-related methods of supporting the rotary and reciprocating parts of a hydraulic motor of the radial piston type whereby hydrostatic and mechanical losses are minimized and the forces imposed on the main rotating parts are most effectively utilized from the standpoint of torque delivery.

Motors of the radial piston type of the present conventional designs are very inefiicient and lack particularly good starting torque characteristics. This is an important deficiency inasmuc as these 5 motors must compete commercially with electro mechanical and electric transmis- (Cl. 121-61) Y and mechanical transmissions wherein the power is more directly applied. [It is an inherent characteristic in hydraulic motors thatthese motors must operate in conjunction with v hydraulic pumps or generators of some sort and; since power hasaiready been transformed once in such pumps or generatorsfurther loss must be minimized as much as'possible in the motor if the total efliciency of the generator-motor unit is to be maintained high enough to compete with transmissions of the other types mentioned.

It is particularly important thatyvhen the pres- I sure power of the generator is weal: the motor used in conjunction therewith must have an efficiency comparable to ordinary geared transmissions or the hydraulic motor is incompetltive from thestandpoint of commercial utility.

In order to accomplish the above objects "and approach as nearly aspossible the ideal characteristics mentioned above, I provide first, an unusually rigid general structure, that is rigid enough to withstand the highest applied pressures and the greatest torque delivery and at the .same time I reduce the frictional losses as much as possible both by the provision of adequate rolling bearings and by the distribution of the load on these bearings such that a balanced condition is maintained at all times. One outstanding characteristic of the motor is the provision of a specially fitted drive shaft for the main rotor, which shaft will be certain to remain in alignment with the rotor under all conditions of increased torque at decreased motor speed as well as less torque at very high'speeds without danger of collapse, deflection or distortion; of the parts and excessive wearon the bearings. I

A further important characteristic is the compact and balanced arrangement of multiple valves. so arranged that at all times the filmspread of the hydraulic operating medium is confined within reasonable limits and uniformly distributed about the working axis of the motor, thus reducing any tendency for the working fluid to distort either the valve carrying member or the rotor pr the driveshaftv operatinglnconiunctiontherewith. I have further provided a definite relationship between the valves and valve ports and the operating cams for the pistons and also the mountings for the pistons by which during all operating con,- ditions both hydrostatic pressures and mechanical forces are in balance.

Other objects and features of the invention will become apparent from the following description relating to the accompanying drawings showing onto the shaft as at 2.

the preferred forms and arrangement of the parts.

In the drawings:

Fig. 1 is a longitudinal central sectional. view of one form of motor;

Fig. 2 is a transverse sectional view taken substantially along the line 1-2 on. Fig. 1, portions of a driving flange, however, being removed for clearneu of illustration of the reactance cams;

Fig. 3 is a transverse sectional view taken along the line H of Fig. 1;

Fig. 4 shows in cross-section (same plane as Fig. 2) the port arrangement of a pintle, which would operate in combination with my special cam profile, shown in Fig. 8; and

Figs. to 10 show diagrammatically the basic curves of my multi-sectional cams and the dead center positions of radial pistons relative to the cams;

Fig. 11 is a left hand end elevation of the cyllnder barrel removed from the associated parts; Fig. 12 is a fragmentary sectional view-taken on a plane indicated by the line l2-l2 in Fig. 11; Fig. 13 is a fragmentary plan view of the barrel with the plungers removed as viewed from the plane 13- in Fig. 11;

Figs. 14 and 15 are sectional views taken on planes indicated by the lines i i-ll and 15-15 respectively, in Fig. 1;

Figs. 16 and 17 are front and end elevations respectively of a plunger and cooperating rollers;

Fig. 18 is a sectional view of a plunger with the thrust means and bearings removed and is take on a plane indicated by the line "-48 in Fig. 1e.

In the illustrated embodiment of Figs. 1 to 3, 1 indicates thedriving shaft which is adapted,

by the use of suitable couplings or connections, to

turn the spindle ofan automatic machine tool or the driving shaft of an automobile or Diesel driven locomotive, for example. The shaft has an enlarged head 3 which is tapered and accurately fitted to.a corresponding tapered bore of the piston carrying rotor 5 supported in the motor casing as will be hereinafter described. The tapered head 3 is'secured against rotation in the tapered bore by means of a feather key 4 and the tapered surfaces aredrawn into stressed engagement with each other by a suitable nut 6 threaded The tapered connection of the shaft to the rotor insures the proper rigid relationship between the shaft and rotor and insures the absolute axial alignment of the shaft and rotor during all working conditions, thereby minimizing any tendency of the shaft to become loose in the rotor and'wobble; it being manifest.

that any wobble, however slight, would, when once set up, render-the entire unit useless in a short time. a

The drive shaft is entirely supported by the rotor 5. The rotor'has hub portions 1 and la at its ends through the medium of which and of interposed bearings I! it is supported by a very rigid casing comprising generally cup-shaped end plates is and 20 having hub portions at 2| and 21 respectively. The end plates 19 and 20 are rigidly connected together by an annular casing member 34, said casing parts being accurately aligned by suitable shoulders and bolted-together in the conventional fashion (bolts not shown). The hub 2| of the casing part I9 is relatively elongated and rigidly carries a stationary pintle 23 of relatively large diameter at its region of support and inwardly therefrom twwarcl the interior of the casing, the pintle pro. acting'into the rotor at one end and having a r duced diameter valve.

portion at 24 in working contact with a complementary valve portion of the central bore 8 in rotor on its bearings l1 with reference to the pintle, this movement being proportionate to the -'operating load on the motor, providing slightly greater clearance when the load is excessive than during lower load output. Considerable clearance is provided as at 1'' between the larger diameter portion of the pintle and the surrounding portion of the rotor as well as at 1' between the radial shoulder which connects the large and small diameter portions of the pintle and the adiacent surface of the rotor. A similar arrangement is disclosed and claimed in my earlier filed copending application, Serial No. 716,451, filed March 20, 1934.

The bearings for the rotor comprise inner bearing rings ll and outer bearing rings ll, the inner rings being tightly fitted, as by press fit onto the rotor hubs 1 and 1a and shouldered thereagainst and the outer bearing rings being accurately seated in suitable shouldered cylindrical surfaces in the casing parts it and 20. The antifriction bearing elements may comprise suitable balls II. The bearings are designed both to resist radial load as well as axial thrust, wherefor the balls are deeply seated in the bearing races and the operating clearance between the balls and race channels is very slight. It will be noted that the balls ii are in radial alignment respectively with the relatively larger portion of the pintle which projects into the clearance spaces 1 and l of the rotor and with the tapered connection between the drive shaft and the rotor. It will be further noted that the main bearing diameters are approximately equal to the overall axial length of the rotor portion embraced by the bearings, this relationship being important to minimize deflection of the axial midportion of the rotor surrounding the reduced valve portion 24 of the pintle wherein the valve'ports are located and to reduce the peripheral speed of the bearings at average speeds of operation of the motor.

Another highly important, relationship, is .between the axial spacing of the bearings and the maximum axial film spread of the working fluid at the pintle valve portion, the actual axial extent of working contact between the valve portion 24 of the pintle and the coacting surface of the rotor being determined by the maximum film spread of the working fluid at maximum pressure of the fluid in driving the motor. It will be noted that the bearings I! are spaced a distance equal .to about twice the length of the cooperating valve portion of the pintle and rotor, in other words, twice the maximum axial film spread at maximum fluid pressure.

In the barrel 5, in the radial plane of the central portion of the reduced valve portion N of the pintle are a suitable number of radial cylindrical bores ii, in which reciprocable pistons ii are carried. The pistons may, for greater economy and rigidity, be made each in one piece and at their outer ends, each has a transverse bore Ila, as better illustrated in Fig. 18,

carrying relatively small needle roller bearings 15 eter portion of the pintle.

- between the pintle valve II and a hardened cross pin l4 embraced and supported by the bearings in true parallelism with the axis of rotation. The oppositely projecting ends of the cross'pins I! carry suitable harden-ed rollers l5 (pressed ,onthecross pins for example) for contact 'with the surfaces of the operating cams which may be formed directly on axially spaced portions 33 and 33' of the casing member 34 in which event the member 34 is of properly selected and treated steel suitable for' use in heavy duty cams. The cam contours, will be hereinafter more fully discussed, but comprise a plurality of separate cam profile sections subject to considerable modification,

but each having both inwardly and outwardly directed surfaces operating as reaction members under the impelling force of the working fluid on the pistons to turn the rotor,-and also to eflect the discharge of the used fluid medium from the motor. 7

As shown in Fig. 3, the fluid medium is preferably fed as through a main inlet 21 in the hub 2i of the plate I! into aman'ifold or distribution groove between the pintle and said hub. From thence the fluid passes through radial openings in the pintle aligned with the distribution groove and into longitudinally extending bores 29 circumferentially spaced within the body of the pintle and extending from the plane of the radial holes 30 to the middle of the reduced diam- Radial inlet ports 30' are formed in the reduced portion of the pintle for communication with the piston cylinders through respective openings 9 in the inner wall of the rotor. The working fluid is thus fed to the cylinders at certain positions of these cylinders about the axis-of the pintle during rotation of the barrel 5. The position and number of the radial ports 30' is determined by the number of cam profllesections or lobes of the cams, that is to say, if there are four profile sections as illustrated in Fig. 2 in equally spaced circumferential relationship then there are fourinlet ports 30' identically spaced about the reduced end of thepintle. This distribution of inlet ports about the axis of the pintle results in that the positive hydrostatic forces of the fluid medium during film spread at the valve are always balanced on radially opposite'zones of the coacting surfaces of the pintle and r'otor and there is, at no time,

any hydrostatic force applied in such manner as to tend to throw the rotor toward one or the other side of its axis as in prior motors of the radial pistontype or to deflect the pintle.

In prior motors of the hydrostatically balanced type, having not more than two complete cam profile sections and therefore only two effective inlet and two effective exhaust ports at the valve, the fllm spread about the entire circumference of the valve cannot, without such clearance between the working surfaces as will too greatly reduce efficiency, be such as to fully lubricate the surfaces and maintain the pressures in hydrostatic balance on diametrically opposite surfaces of the valve at the same time. My arrangement of valve ports, close together, accomplishes this, since at all times minimized positive pressure fluid film spread will be present as much as necessary for proper lubrication notwithstanding unusually close working clearance portion 24 and the wall of the rotor bore 5'.

It should be further noted that by the arrangement of ports close together, the circumferential film spread during operation of the .charged through the mizes both skin friction and the discharging fluid.

motor is minimized between the working surfaces of therotor and pintle (valve portion 24 and this is one of the outstanding advantages of my arrangement of cam profile sections and valve ports, wherein the valve ports are sufilciently close to one another to effect instantaneous fiuid film spread on the entire periphery of the valve portion 24 of the pintle to immediately effect hydrostatic balance and thus prevent the.locking effect of unbalanced hydrostatic pressure which would expand with various spread on the surface according to the smoothness or roughness of the adjacenthydraulically fitted surfaces at the bottommot'or speed.

The working fiuid after operating to move the pistons radially outwardly is thereafter discylinder ports 9 into short valve ports l,..and 82, which communicate didectly with'an enlarged central bore 28 in the pintle, which bore 28 has one end terminating in radial alignment with the said short exhaust Ports and the other end open asjat 26, as illustrated in Figs. 2 and 1 respectively, for connection with suitable piping to lead the fluid medium from the motor. This valve and duct arrangement provides for the free dischargeof the operating fiuid medium from the motor cylinders without imposing any appreciable dead direct the discharge openings and ducts for exhaust operating fluid are, theless drag or dead load will be imposed on the motor in ejecting the fiuid. The present solution practically miniviscous friction of In order that the wear on the pistons will be minimized I provide, on the rotor La radial driving flange 8 which extends outwardly from the rotor I in the plane of the cylinders, and the cylinder bores are extended outwardly through said flange, and form guideways 8a therein and provide slots 8' for the crosspins II, as better illustrated in Figs. 11 to 15 inclusive. It will be seen that by this arrangement the pistons ,are fully supported in a uniform manner throughout their entire length both fore and aft in the direction of rotation of the rotors and there can be no possibility of the peripheral forces imposed on the cross pins by the cams enlarging or distorting the cylinder bores unnecessarily or wearing theadjacent surfaces of the pistons or in any manner deflecting the pistons transversely of their axes.

Each cam is divided into pressure and exhaust portions, the circumferential extent of a given cam portion which acts on a piston to cause it to complete one pressure and one exhaust stroke, beingherein referred to as a cam profile section. Otherwise stated each cam comprises a connected series of convex and concave profiles.

Referring now to the cam contours, the cam profile sections, illustrated in Fig. 2, each comprofile sections are shown number of complete profile sections including both inwardly and outwardly directed surfaces than two.

The basic contours of the inwardly and outwardly directed cam surfaces are illustrated in Figs. 5 to 10, wherein the power applying and fluid discharge portions of the individual cam as meeting abruptly; however, it is understood that the meeting portions of the curves would be rounded more or less, depending upon the starting and running torque and speed requirements of the motor. In Fig. 5, I represents the rotor and the radial lines of symmetry "-42 denote the dead center positions of the pistons'on the cams, as well as the separation lines of the cam sections.

Figs. 6 and 8 illustrate that an uneven number of profile sections may be provided with, of course, the proper numbefoi' inlet and exhaust ports in the valvewhich number is twice the number of cam profile sections. Fig. 6 shows a quinquefoil shows a trefoil arrangement. The radial lines of symmetry-denoting the dead center positions of the pistons on the cams are indicated from 43 to l! on Fig. 6 and-the cams of Fig. 8 are similarly divided. Each of the cams has twice as many dead center positions as its name calls for; a treposition foil six, a quatrafoil eight etc.

The cam arrangement according to Fig. 7 is characteristically different from the previously described cam forms in that there is a very short, relatively convex, cam surface only. The dead center positions on the cams are partially indicated by radial lines "to I. This form of cam is particularly adapted to motors designed for low speed and great torque. A motor with the piston arrangement of Fig. 2 and the cam arrangement of Fig. 7 would operate with very little vibration since the starting torque would be very gradual while the piston is moved from the line ll counterclockwise toward the line 54, it being obvious that with the cross pin rollers at the position 5, the force component tending to drive the rotor in the direction of the arrow would be very small but increasing proportionately with the rotational angle as 54. Such a motor would approach the torque characteristics of a series electro-motor with the starting torque increasing directly with the load. In like manner the starting torque characteristics of the motor with any of the other cam arrangements may be varied by the modification of the cam contours. For example, the cam arrangement of Fig. 8 would approach more nearly a triangle with the rounded spices and the arrangement of Fig. 6 would approach more nearly a pentagon with rounded spices.

In any case the driving torque will be exerted by the tangential component of each radial piston load. Since the pistons are all radial the piston load will be radial and this radial load will have a projected component into and on the geometrical tangent of the cam curve in that particular point of the cam path in which the center line of the piston intersects the cam profile curve. The geometrical tangents of the cam curves will be perpendicular to the cam curve radii which pass through the center of curvature of .each concave and convex portion of the cam.

arrangement of cams and Fig. 8.

the rollers approach the Thus in analyzing the torque characteristics of the motor with the different cams the actual tangential force and its torque may be graphically determined and illustrated. The torque naturally will be the product of the tangential force and the torque radius (instantaneous).

It has beenfdemonstrated that with the arrangement of four cam profile sections and six pistons, as shown in Fig. 2, whenever the pistons are moving outwardly under the positive force of the fluid medium, the same mechanical force is applied to the pistons on opposite sides of the center of rotation of the rotor and that the'hydrostatic forces which are set up by the film spread at the valve are balanced on opposite sides of the same center of rotation. These conditions obtain from the inward dead center positions of the pistons until the pistons are in their outermost dead center positions, (not shown) and thereafter any mechanical forces imposed by the action and 7. This identical ratio cannot be maintained for the cam arrangement of Fig. 8 e. g. but nevertheless by providing 12 pistons, the operative characteristics of the motor and balanced condition of forces will be the same as heretofore described. In the Fig. 6 arrangement of cams one manner in which results comparable to those obtained by the Fig. 2 arrangement could be obtained would be by providing 20 pistons, in which event theabove described balance-and starting characteristics would be fulfilled,

It may be found impractical to provide so great a number of pistons, as required for Figs. 6 or 8 e. g., in a single series, wherefore I may divide the required numbers into appropriate groups and arrange the groups in axially offset relationship with reference to the pintle. For instance for the three section cam of Fig. 8 I may provide two groups of 6 pistons each and with a valve port arrangement according to Fig. 4 for each group of pistons (see also dotted line indication of valve 58, 2, iii, 6! and represent one maining pistons of the same group are in exhaust positions and all the pistons of the other group are in dead center positions.

It will be clear from Fig. 8, assuming the 'valve ports of the valve for as shown diagrammatically on this flgure, that the rotor will turn pistons 58', U0 and 62 continuing to turn the rotor until the cylinder ports of the pistons 58, N and M are aligned with the respective inlet ports Iii. The operation just described is the same notwithstanding whether all twelve pistons are in one plane or separated into groups sav of both groups of pistons are in a clockwise direction, the it 2,101,899 three or of six pistonseach in different planes along the rotor axis.

Referring to Figs.'9 and 10, these iliustratethe basic characteristics of a two-lobe cam arrangement embodying the principles of the present invention. Fig. 9 shows the theoretical generation of the cam surfaces, the lines 10 and H indicating minor and major-axes of the cam and the points 12 disposed on the axis ll equal distances from the axis 10 being the points of generation of the cam arcs 13. These arcs I3 may intersect as at 14 forming a dead centerpoint of unstable equilibrium for the pistons in their innermost posi-, tions on opposite sides of the axis H, or, alternatively at Hie region of the "intersection of. the arcs I3, the corners may be rounded as-on arcs 14' for smoother operationofthe"motor.-

Fig. 10 shows [the preferred: arrangement of ports in the pintie and an exemplary arrangement of pistons designated 15 to inclusive. A

four-port valve is recommended having intakeports 30 and exhaust ports 32 communicating with axial passages in thepintle as previously described. -Also the port arrangement provides bridge portions at 32' between the inlet and exhaust ports. It will be seen that inthe operation of a motor designed with this cam and the port arrangement and relationship thereto shown, the motor will turn in a clockwise direction, pistons 16 and I! being in approximately maximum torque positions, pistons 11 and 80 being in exhaust positions and pistons 15 and 18 being momentarily on dead center at the juncture of the cam surfaces defined by the arcs l3.

It will be seen that in all the above described arrangements of earns, the cam surfaces may be machined with standard machine tools and, after machining, be finished on a circular grinder and the rounded portions of the cams, where the arcs l3 defining the major portions of the cam surfaces meet, may be similarly simply formed. Such simple manufacturing methods would be impossible on prior forms of cams, say elliptical cams.

It will be noticed that by virtue of my novel cam profile arrangement, whereby there are a circular series of cam lobes in the basic curves, the rotating pistons of the motor will balance each other substantially in regard to their centrifugal and tangential mass forces also. This dynamic balance is an additional feature of my novel improved hydro-motor, which is lacking in the motors of the prior art. The combined dynamic and hydrostatic balance of my motor allows a greater maximum speed and more uniform rotation, together with less fluctuation of the driving power of the motor, than it is obtainable with devices of similar nature of today. Since the upper speed ratio of my motor, when it is used in combination with a variable delivery pump, is substantially high, compared to the lower limit of the speed, I claim that due to the .substantial dynamic balance of the pistons at in the. case of unbalanced piston arrangements;

I claim:

1. In a high pressure rotary radial piston hydraulic apparatus, a casing, a barrel comprising a rigid cylindrical portion having coaxial hub portions near the ends respectively, Sets of anti-friction bearings mounted in the casing and on said hub portions respectively for supporting the barrel for rotation about its axis, said barrel having a set of circumferentially spaced radial cylinders disposed. between the planes of the sets of bearings, a radial flange on said barrel in the plane of said cylinders and spaced axially of the barrel inwardly from the sets of bearings, said flange having radialguideways respectively aligned with the cylinders, plungers in said cylinders having end portions extending radially outwardly therefrom and lying in the associated guideways respectively, said plunger end portions being coextensive axially of, the barrel with the flange, crosspins mounted in the said end porside walls extending radially of the flange alongside the radial faces thereof and in substantially abutting relationship thereto, said tracks engaging the plunger rollers for relatively rolling engagement therewith for effecting rotation of the barrel, and valve means invalving cooperation with said cylinders.

2. In a high pressure rotary radial piston hydraulic motor, a casing, a barrel comprising a rigid cylindrical portion having coaxial hub per-- tions at the ends respectively, sets of anti-friction bearings mounted in the casing and on said hub portions respectively for supporting the barrel for rotation about its axis, said barrel having a set of circumferentially spaced radial cylinders disposed between the planes of the sets of bearings, a radial flange on said barrel in the plane of said cylinders and spaced axially of the barrel from the sets of bearings, said flange having flat parallel radial side walls and having radial guideways. respectively aligned with the cylinders,

plungers in said cylinders having end portions eral faces lying in the planes of the radial side walls of the flange respectively, crosspins mounted in the said end portions of the plungers and extending parallel to the barrel axis and beyond the said lateral faces, rollers on the ends of each crosspin and on their inner faces in substantially face to face abutting relation to the flat lateral faces and radial side walls of the flange'and associated plunger end portion, axiallyspaced parallel cam tracks surrounding the barrel between the planes of the bearings and having flat radial side walls extending radially of the flange alongside the radial side walls thereof and in substantially abutting relationship thereto, said tracks engaging the plunger rollers for relatively rolling engagement therewith for effecting rotation of the barrel, and valve means in valving cooperation with said cylinders.

3. In a high pressure rotary radial piston hydraulic motor, a casing, a barrel comprising a rigid cylindrical portion having coaxial hub portions near the ends respectively, sets of anti-friction bearings mounted in the casing and on said hub portions respectively for supporting the barrel for rotation about its axis, said barrel havcylinders and spaced axially of the barrel from' the sets of bearings, said flange having flat radial side walls and having radial guideways respectively aligned with the cylinders, plungers in said cylinders having end portions extending radially outwardly therefrom and lying in the associated guideways respectively, said plunger end portions .being coextensive axially of the barrel with the flange, thrust means carried by the said end portions of the plungers and extending parallel to the barrel axis and beyond both lateral limits of the flange, axially spaced parallelreactanee means surrounding the barrel between the planes of the bearings and having flat radial side walls extending radially of the flange alongside the radial faces thereof and in substantially abutting relationship thereto, said reactance means engaging the thrust means contiguous to the lateral faces of the flange for effecting rotation of the barrel, the pitch diameter of said sets of bearings deflning a cylindrical surface of substane tially the same diameter as the cylindrical portion of the barrel, and valve means fltting in said barrel bore between the planes of the sets of bearings and in valving cooperation with said cylinders and terminating between the planes of said bearings."

,4. In a high pressure rotary radial piston hydraulic motor, a casing, a barrel comprising a rigid cylindrical portion having coaxial hub portions at the ends respectively, sets of anti-friction bearings mounted in the casing and on said hub portions respectively for supporting the barrel for rotation about its axis, said barrel having a set of circumferentially spaced radial cylinders disposed between the planes of the sets of bearings, a radial flange on said barrel in the plane of said cylinders and spaced axially of the barrel from-the sets of bearings, said flange having flat parallel radial side walls and having radial guideways respectively aligned with the cylinders, plungers in said cylinders having end portions extending radially outwardly therefrom and lying in the associated guideways respectively, said plunger end portions being coextensive axially of the barrel with the flange and having flat lateral faces lying in the planes of the radial side walls of the flange respectively, each plunger end portion having a bore extending parallel to the barrel axis, crosspins in said bores of the said end portions of the plungers and extending parallel to the barrel axis and beyond the said lateral faces, rollers on the ends of each crosspin and on their inner faces in substantially face to face abutting relation to the flat lateral faces of the associated plunger end portions respectively and the radial side walls of the flange, sets of capillary needle rollers for mounting the crosspins in said bores respectively, said needle rollers being substantially coextensive axially of the associated crosspins with the portion of the crosspin lying between the associated rollers, axially spaced parallel cam tracks surrounding the barrel between the planes of the bearings and having flat radial sidewalls extending radially of the flange alongside the radial side walls thereof and in substantially abutting relationship thereto,

"said tracks engaging the plunger rollers forrelarotation of the barrel, and valve means in valving cooperation with said cylinders.

5. In a radial piston pump or motor, a casing,

a rotatable barrel, a reactance means, piston and cylinder assemblies carried by the barrel and cooperating with the reactance means, said barrel having an axial valve bore and ports operatively communicating the assemblies respectively with said valve bore, a pintle mounted in the casing and having a valve portionextending into the axial bore of the barrel'and fltting said axial valve bore, said pintle valve portion having a plurality of circumferentially spaced high pressure fluid ports and low pressure fluid ports successively communicable with the ports of said assemblies as the barrel rotates, a plurality of circumferentially spaced ducts extending within and longitudinally of the pintle and connected respectively to different ports of one group of said pintle ports and each being radially spaced throughout its length from the pintle axis, and a single duct of larger cross sectional area than said spaced ducts extending substantially axially within said pintle and being connected to all of the ports of the other group of pintle ports and extending alongside the first mentioned ducts.

6. In a high pressure, rotary, radial piston hydraulic motor, a barrel comprising a rigid cylindrical portion having circumferentially spaced radial cylinders therein, valve means for the cylinders, anti-friction bearing means at oppo-r site sides of the plane of the cylinders and supporting the barrel for rotation about its axis, a radial flange on said cylindrical portion in the plane of the cylinders, and having radial guideways respectively aligned with the cylinders, plungers in the cylinders, each plunger having a portion reciprocable in and guided by an associated guideway and substantially coextensive axially of the barrel with the flange, crosspins anti-frictionally mounted in the said portions of the plungers and extending parallel to the barrel axis beyond the lateral limits of the flange, rollers on the respective ends of each crosspin and having their inner faces in substantially face to face engaging the plunger rollers substantially contiguous to the lateral faces of the flange for relative rolling cooperation therewith for effecting rotation of the barrel.

7. In a high pressure, rotary, radial piston hydraulic motor, a barrel comprising a rigid cylindrical portion having circumferentially spaced radial cylinders therein, valve means for the cylinders, anti-friction bearing means at opposite sides of the plane of the cylinders and supporting the barrel for rotation about its axis, a radial flange on said cylindrical portion in the plane of the cylinders, and having radial guideways;respectively aligned with the cylinders, plungers in the cylinders, each plunger having a portion reciprocable in and guided by an associated guideway and substantially coextensive axially ofithe barrel with the flange, crosspin means mounted in said portions of the plungers and extending parallel to the barrel axis beyond'thelatera'l'limits of the flange, sets of capillary needle rollers for rotatably mounting the crosspinmeans in the associated portions, said needle rollers being substantiallycoextensive axially with the associated ting relationship thereto, said reactance means plunger portions, axially spaced parallel reactance means surrounding the barrel and having side walls extending radially of the flange alongside and in substantially abutting relation to the radial faces thereof, and said reactance means operatively engaging the extending portions of the ELEK K. IBENEDEK. 

